The present invention relates to a vehicle""s automatic transmission. In particular, the present invention relates to a hydraulic control system and method for automatic transmissions capable of improving gasoline mileage and shift quality by adjusting hydraulic pressure applied to friction elements of the automatic transmission.
Typically, a transmission controller of an automatic transmission shifts between gears by controlling a plurality of solenoid valves installed on hydraulic lines, based on driving conditions of the vehicle, such as an operational speed of the vehicle, a degree of throttle opening, or the like. That is, if a shift lever is manipulated to a target position, a manual valve operates to change or convert its ports such that hydraulic pressure from an oil pump is supplied to appropriate operating elements of the gear shift mechanism, according to a duty control of the solenoid valves under control of the transmission controller.
When the transmission operated by the shift lever is positioned at a desired range, the hydraulic pressure is applied to some friction elements and released from other friction elements so as to establish a predetermined gear ratio. Accordingly, the transmission performance depends on the timing of applying and releasing hydraulic pressure to and from the predetermined friction elements for the target gear ratio.
In the case of an electrical transmission control system in which gear shifting is performed by applying hydraulic pressure to the friction elements, the pressure level applied to the friction elements is a stable standard line pressure. The hydraulic pressure is supplied from an oil pump that is mechanically connected to the engine so as to operate according to the engine operation, and the oil pump should be designed to sufficiently supply the hydraulic pressure in various ranges of engine rpm, starting about 700 to 800 rpm, and to maintain the hydraulic pressure so as to be failsafe even under the worst conditions.
The line pressure is set to satisfy various driving conditions such that if the line pressure is adjusted according to driving conditions it is possible to reduce the power wastage caused by the oil pump, resulting in improved gasoline mileage.
However, in the case of an automatic transmission performing a gear shift using a clutch to clutch method, the standard line pressure may change while adjusting the line pressure such that the friction elements from which the hydraulic pressure is released and the friction elements to which the hydraulic pressure is applied are simultaneously controlled. This is unlike the clutch to one-way clutch method, resulting in difficulty of line pressure control and degradation of shift control stability.
Furthermore, in conventional transmission control, it is impossible to estimate the time taken for the line pressure to reach the predetermined pressure level in the case of shift delay control, during which the line pressure increases to a predetermined pressure level before shifting. Also, it is required to determine friction coefficients of the friction elements of the transmission, deviation of the hydraulic pressure, and durability for the control of the line pressure in an in-gear state in which a certain shift range is synchronized without shifting gears.
In a preferred embodiment of the present invention, the hydraulic control system for an automotive automatic transmission comprises a driving condition detection unit, a transmission control unit, and a driving unit. The driving condition detection unit detects driving conditions of a vehicle. The transmission control unit performs variable line pressure control using a minimum line pressure and a line-pressure-decreasing gradient calculated based on driving condition data detected by the driving condition detection unit when the driving conditions satisfy variable line pressure control entrance conditions. The driving unit adjusts a duty ratio of line pressure applied to friction elements according to a line pressure control signal generated by the transmission control unit.
It is preferable that the driving condition detection unit comprises: an engine rpm sensor for detecting engine rpm of the vehicle; a throttle-opening sensor for detecting throttle opening degree; a turbine rpm sensor for detecting turbine rpm of a torque converter of the vehicle; a driving shaft rpm sensor for detecting rpm of a driving shaft of the vehicle; a hydraulic fluid temperature sensor for detecting hydraulic fluid temperature of the transmission; an atmospheric pressure sensor for detecting atmospheric pressure of an area where the vehicle is running; and an inhibit-on switch for detecting a position of a shift lever.
Preferably, the transmission control unit calculates a compensation value for variable line pressure control, based on the throttle opening degree.
It is preferable that the transmission control unit controls such that the variable line pressure control duty increases according to the change of the throttle opening degree and then is eliminated when engine power is normalized. It is further preferable that the transmission control unit calculates a line pressure control compensation value and hydraulic characteristics according to the hydraulic fluid temperature and engine rpm, and applies the calculation to the variable line pressure control duty. Preferably, the transmission control unit calculates a compensation value for a clutch friction coefficient according to a deviation and durability of the transmission and reflects the compensation value to a variable line pressure control duty.
It is preferable that the transmission control unit learns a minimum required line pressure by detecting an in-gear slip.
Preferably, the transmission control unit sets a delay time for increasing a shift begin line pressure to 100% for control stability.
It is preferable that the transmission control unit sets a minimum delay time by estimating a line pressure at a shift begin point without a hydraulic pressure sensor and a time taken for the line pressure reach to 100%.
It is preferable that the transmission control unit sets a minimum shift delay time in a slow kick-down.
Preferably, the transmission control unit sets a minimum shift delay time during lift-foot-up.
In another preferred embodiment of the present invention, the hydraulic control method comprises: determining whether or not driving conditions detected in a non-variable line pressure control mode satisfy variable line pressure control entrance conditions; determining whether or not gears are in an in-gear state by calculating a gear ratio using an engine rpm and turbine rpm when the driving conditions satisfy the variable line pressure control entrance conditions; entering a variable line pressure control mode in an in-gear state, and performing the variable line pressure control by calculating a minimum line pressure and variable line pressure gradient; determining whether or not a present line pressure is less than the calculated minimum line pressure; performing a normal line pressure control when the present line pressure is less than the minimum line pressure and then determining whether a gear shift starts or a damper clutch is in a direct coupled state; performing a gear shift to a target gear ratio after a predetermined time from a point when a line pressure control duty reaches 100% if it is determined that the gear shift starts or the damper clutch is in the direct coupled state in normal line pressure control; and performing variable line pressure control according to the driving conditions by entering the variable line pressure control mode for determining another shift begin point after delaying for a predetermined time if the gear shift to the target gear ratio is completed.
Preferably, the variable line pressure control mode entrance conditions include: the present hydraulic fluid temperature is between a preset lowest limit minimum value and lowest limit maximum value; a CAN (Controller Area Network) communication line providing an interface for various control data and detection signals is not broken down; an inhibit-on switch, hydraulic sensor, and line pressure solenoid valve are normal; a present detected atmospheric pressure value is less than a threshold atmospheric pressure value for determining whether or not the vehicle is running at a high altitude; a shift lever is positioned at one the D, 4, 3, and 2 ranges, or 2, 3, 4, and 5 ranges in sports mode; a throttle opening degree (TPS) is less than a preset threshold opening degree with a compensation value added thereto; an engine rpm is less than a threshold engine rpm; and engine rpm detection and turbine rpm detection are performed normally.
It is preferable that driving condition data do not satisfy the variable line pressure control mode entrance conditions, a line pressure control mode returns to a non-variable line pressure control mode.
Preferably, a line pressure control mode returns to a non-variable line pressure control mode if an in-gear state is not detected.
It is preferable that the in-gear state is a state where certain gears are engaged for a predetermined gear ratio according to driving conditions.
It is also preferable that the variable line pressure control is performed in such a way that if it is the first in-gear state after reset of the battery, the transmission control unit decreases the line pressure from 100% by the gradient per cycle and stores count values in respective learning areas.
Preferably, in a case of performing variable line press control according to respective learning areas, the variable line pressure control is not performed at 100% of line pressure but at the point (A+[(Dvfs)min]) where the line pressure is increased by as much as a predetermined percentage at the minimum required line pressure duty.
It is preferable that the minimum line pressure calculation includes: detecting an engine brake torque (TB); detecting a turbine torque (TT); detecting a required line pressure (PL); detecting a standard value (D_BASE) of the required line pressure from a standard duty of the line pressure; and calculating the minimum line pressure by adding various compensating coefficients to the standard value (D_BASE) of the required line pressure.
It is preferable that if an in-gear slip is detected while the variable line pressure control controls to the minimum line pressure, a new minimum line pressure is learned according to the driving conditions and vehicle""s durability, and then the minimum line pressure is reflected to the variable line pressure control.
It is preferable if another gear shift begins or a damper clutch is in a direct coupled state, the line pressure duty increases to 100%, and then the gear shift to the target gear ratio is performed after a predetermined time delay.
It is preferable that if a throttle opening degree is changed in the variable line pressure control, the line pressure is compensated according to the change of the throttle opening degree.
Preferably, if an in-gear slip is detected in a normal line pressure control procedure, a minimum line pressure is learned according to the driving conditions and vehicle""s durability and applied for the line pressure control.
It is preferable that the delay time is set, in a map table, based on the line pressure at a shift begin point in a power-on up-shift condition, based on the line pressure at a point prior to a predetermined period from the shift begin point in a power-off up-shift condition, based on a value obtained by subtracting a slow kick-down compensation value (Tsk) from a map value (Tdo) set at a point prior to a predetermined period from the shift begin point at a power-on down-shift condition, and based on the line pressure at the shift begin point in the power-off down-shift condition.
It is preferable that the engine brake torque (TB) is calculated using a maximum engine torque (TQ_STND), a compensation vale (TOI_hex) obtained based on the driving conditions such as an intake air amount, temperature of the intake air, fuel injection amount, ignition point, and the like, and a torque loss caused by engine friction, according to the following equation:
TB=TQxe2x80x94STND*(TQI_hexxe2x88x92TQFR_hex)/255/9.8
It is preferable that the turbine torque (TT) is calculated using a torque ratio (tr) of a torque converter obtained according to a ratio (Nt/Ne) of an engine rpm (Ne) and turbine rpm (Nt) in a map (TTRQRTP) of a ratio of the engine torque and torque converter torque, according to the following equation:
TT=TB*tr
It is preferable that in the 4 range automatic transmission, the required line pressure (PL) in the state where the damper clutch (D/C) is directly coupled is calculated as in the following equation:
PL=turbine torque coefficient (XVFxe2x80x94PTDC)xc3x97safe factor (XVFxe2x80x94SF)xc3x97turbine torque (TT);
it is calculated as in the following equation 14 when the damper clutch (D/C) is not in the direct coupled state:
1st gearxe2x88x92PL=XVFxe2x80x94PTAxc3x97XVFxe2x80x94SFxc3x97TT+XVFxe2x80x94OFB
2nd gearxe2x88x92PL=XVFxe2x80x94PTAxc3x97XVFxe2x80x94SFxc3x97TT+XVFxe2x80x94OFB
3rd gearxe2x88x92PL=XVFxe2x80x94PTAxc3x97XVFxe2x80x94SFxc3x97TT+XVFxe2x80x94OFB
4th gearxe2x88x92PL=XVFxe2x80x94PTAxc3x97XVFxe2x80x94SFxc3x97TT+XVFxe2x80x94OFB;
and in a 5 range automatic transmission, the required line pressure (PL) in the direct coupled state of the damper clutch (D/C) is calculated as in the following equation:
2nd gearxe2x88x92PL=XVFxe2x80x94PTDCAxc3x97XVFxe2x80x94SFxc3x97TT+XVFxe2x80x94OFB
4th gearxe2x88x92PL=XVFxe2x80x94PTDCAxc3x97XVFxe2x80x94SFxc3x97TT+XVFxe2x80x94OFB
5th gearxe2x88x92PL=XVF_PTDCAxc3x97XVFxe2x80x94SFxc3x97TT+XVF_OFB;
when the damper clutch (D/C) is in the direct coupled state at any of the 1 and 3 range, the required line pressure (PL) is calculated as in the following equation:
PL=XVFxe2x80x94PTDCxc3x97XVFxe2x80x94SFxc3x97TT;
and at the normal ranges where the damper clutch (D/C) is not in the direct coupled state, the required line pressure (PL) is calculated as in the following equation:
1st gearxe2x88x92PL=XVFxe2x80x94PTAxc3x97XVF_SFxc3x97TT+XVFxe2x80x94OFB
2nd gearxe2x88x92PL=XVFxe2x80x94PTAxc3x97XVFxe2x80x94SFxc3x97TT+XVFxe2x80x94OFB
3rd gearxe2x88x92PL=XVFxe2x80x94PTAxc3x97XVFxe2x80x94SFxc3x97TT+XVFxe2x80x94OFB
4th gearxe2x88x92PL=XVFxe2x80x94PTAxc3x97XVFxe2x80x94SFxc3x97TT+XVFxe2x80x94OFB
xe2x80x835th gearxe2x88x92PL=XVFxe2x80x94PTAxc3x97XVFxe2x80x94SFxc3x97TT+XVFxe2x80x94OFB;
wherein when the required line pressure (PL) is less than the preset minimum line pressure [3.2(XVF_PLMIN)], the required line pressure (PL) is set equal to the minimum line pressure [3.2(XVF_PLMIN)], XVF_PTA is a turbine torque coefficient for calculating the required line pressure (PL) at the corresponding range, XVF_SF is a safety factor which is about 1.2, TT is a turbine torque, and XVF_OFB is an offset value for calculating the required line pressure (PL) at the corresponding range.
It is preferable that minimum line pressure [(Dvfs)min]] is calculated as in the following equation:
(Dvfs)min=(Dxe2x80x94BASE+Dxe2x80x94L)xc3x97Cxe2x80x94TEMPxc3x97Cxe2x80x94NE+Dxe2x80x94TH,
where D_L is a learned value of the line pressure duty, C_TEMP is a hydraulic fluid temperature compensation value, C_NE is an engine rpm compensation value, and D_TH is a throttle-opening compensation value.
It is preferable that the in-gear slip is determined: when the value obtained by subtracting the turbine rpm (Nt) from the engine rpm (Ne) is greater than a predetermined first threshold rpm, or an absolute value obtained by subtracting the present turbine rpm (Nti) from the previous turbine rpm (Nt) is greater than a predetermined second threshold rpm, while the damper clutch is in direct coupled state; or when the absolute value obtained by subtracting the present turbine rpm (Nti) from the previous turbine rpm (Nt) is greater than a third threshold rpm, while the damper clutch is not completely coupled.
It is preferable that the slow kick-down compensation value is calculated as in the following equation:
Tsk=sumxcex94Dxe2x80x94VFSxc3x97Csk
where sumxcex94D_VFS is a rate of change of the line pressure control duty value (D_VFS) between a point prior to a predetermined period from the SD and the SD, sumxcex94D_VFS can be expressed as in the following:
sumxcex94Dxe2x80x94VFS=xcex94Dxe2x80x94VFS(ixe2x88x92x)+xcex94(Dxe2x80x94VFS(ixe2x88x92x+1)+ . . . +xcex94(Dxe2x80x94VFS(ixe2x88x922)+xcex94(Dxe2x80x94VFS(ixe2x88x921))
where x is the slow kick-down compensation value, Csk is a compensation measurement which is expressed in unit of ms/%, and xe2x96xa1D_VFS(j)=D_VFS(ixe2x88x92x)xe2x88x92D_VFS(j), D_VFD(ixe2x88x92x) is the line pressure control duty value (D_VFS) at the point prior to a certain period from the SD and expressed in the unit of %.
Preferably, the throttle-opening compensation value (D_TH) is calculated as in the following equation:
Dxe2x80x94TH=sum[Dth(ixe2x88x92x)+Dth(ixe2x88x92x+1)+Dth(ixe2x88x92x+2)+ . . . +Dth(ixe2x88x922)+Dth(ixe2x88x921)]
where Dth(i)=(dVth/dt(i)) * Cth is expressed in unit of %, dVth/dt(i) is a change rate of TPS [V/s](calculated per cycle); however, in case of dVth/dt(i)xe2x89xa60, dVth/dt(i) is set to 0, Cth is compensation factor [%N/s], x is a compensation time (XVF_THLDTH[ms]/16 ms) according to the change of the throttle opening degree.